The present invention relates to guided rotary sliding vane machinery, such as compressors, in which the radial motion of the vanes is controlled to obtain non-contact sealing between the vane tips and the interior stator casing sidewall as a result of the cooperation of opposing vane extensions that engage cooperative circular radial guides that are located on both ends of the machine.
Conventional and elementary sliding rotary vane machines are, on one hand, distinguished from virtually all other fluid displacement machines in their remarkable simplicity. On the other hand, such machines exhibit relatively poor operating efficiency. This poor energy efficiency is rooted directly in machine friction, both mechanical and gas dynamic. As is well known, the predominant source of mechanical friction in conventional production contact vane rotary machines occurs at the intense rubbing interface of the tip of the sliding vane and inner contour of the stator wall. Furthermore, governing the motion of the vane by the stator wall contour necessarily and greatly inhibits the area through which gas can enter or exit the machine because ports in the stator housing diminish the vane tip contact area that is already at a premium in such machines. Minimizing flow part area in favor of vane tip contact area results in increased fluid flow pressure losses in the inlet and outlet port regions of such type machines.
Over the years, many means have been proposed to eliminate guiding the radial motion of the vanes through the direct action of the vane tips rubbing along the inside casing or stator wall. In many previous endeavors to grapple with this mechanical problem, attention has been focused upon the use of wheels and rollers pinned to the sides of the vanes wherein these rollers follow inside a circular or non-circular cam track of the appropriate configuration. The cooperation of the rollers in the roller guide track then produces a means of dictating the radial location of the vane which is pinned to the roller follower and hence determines the position of the tip of the vane.
As attractive as this approach first appears to be, rollers wheels contain an overwhelming flaw. Specifically, they cannot provide positive bi-axial radial motion without having to reverse their rotational direction. Thus, vanes constrained by rollers can accommodate geometric displacement in only an outward or inward direction at any one time.
As an example, if the roller has been in contact with one side of the track, and because of this, it has been turning in the clockwise direction, and then should the roller be caused to come in operative contact with the other side of the guide, the roller will be turning in what may be regarded as the wrong direction. As a consequence, each roller undergoes skidding inside the track until it is stopped. Then each vane roller must reverse rotational direction and accelerate to the speed which will match the motion dictated by the other side of the roller guide. Because, in practice, vane machines generally require both positive inward and outward vane motion, rollers become impractical or non-functional in real machines where both motions are often demanded.
Other innovators have taught the use of sliding arc segment tethers in place of vane rollers. In such prior art instances, the arc segment tethers are captured within a circular annular groove that may or may not be rotatable. The arc segment vane tether has the outstanding and fundamentally important advantage of being able to deliver both positive inward and outward radial motion to the vane simultaneously. However, conventional vane motion control techniques use arc segment vane tethers which entailed considerable mechanical friction that arises from the sliding of the arc tether surfaces against the circular annular guides, whether or not the guides themselves are rotatable.
Further, and of fundamental importance, previous workers have failed to provide teachings of the specific contour that the internal casing profile must take on in order for the vane tips to mate closely in a non-contact but sealing relationship with this casing contour profile. Earlier innovators have either simply erred and believed that the proper casing contour was "circular", or circumvented the fundamental issue by characterizing the shape of the internal casing contour with words such as "substantially circular" and not teach operatively just how this fundamentally important shape is determined.
As an example of the foregoing problems, reference is made to U.S. Pat. No. 2,469,510 which shows a rotary vane machine which uses a static extension from the vane, the outside of which bears directly on the inner race of a standard ball bearing to limit outward radial motion. Inward radial motion is limited by the bearing contact of the underside of the vane extension against the outer race of a second conventional ball bearing. Springs shown in FIG. 5 are added to provide a positive outward radial vane bias. As the rotor rotates (assuming, for example, that the centrifugal forces are adequate or that springs are used) the static vane extensions engage the inner race of the outer standard ball bearing. Because the rotational speed of the vane extensions is constantly changing due to the eccentricity of the rotor with respect to the stator (this eccentricity is necessary, of course, for the volumetric changes required), there results a significant component of direct frictional sliding of the vane extension surface with respect to the rotating races of the respective ball bearings.
The inherent enforced sliding component of the vane extensions is about one-quarter of the pure sliding that would take place if the races did not rotate, i.e. a static annulus. While a three-fourths reduction in sliding appears to be significant, the remaining portion is still so large as the be impractical from the fundamental standpoints of wear, reliability, and friction. Significantly, the remaining 25% component of pure sliding is a minimum case. This minimum will be reached only if the rotating races, propelled by the friction arising among all the vane extensions rubbing on the bearing races, are identical to the rotor speed. Of course, this condition can never arise in actual practice because the friction between the vane extension and the rotating surface of the ball bearing constantly varies as the rotor turns. For example, when the vane is most extended, the centrifugal force generated by the motion of the vane and its mass is significantly higher than when it is at the least extended position. This situation is aggravated by the fact that not only does the peak rubbing force arise when the vane is in its outermost position, but the friction moment arm of the vane extension is also largest in that position. Therefore, the vanes that are more extended will tend to accelerate the rotation of the rotating race of the ball bearings at the expense of the vanes which are less extended. This real effect is estimated to be about double the rubbing friction. Therefore, an actual machine would have to contend with a much larger amount of friction and wear on the order of half that experienced by a conventional unconstrained vane machine. A commercial compressor of this type cannot be built because of this inherent problem.
U.S. Pat. No. 4,958,995 is another example of conventional vane pumps with, however, an objective to provide anti-friction vane control. For instance, FIGS. 27 to 29 show a large diameter ball bearing-mounted rigid rotating annular-grooved plate that performs the same function as the rigid rotating annular ball bearings, but requires that the vane extensions be placed nearer the tips of the vanes. This large diameter plate actually rotates with the rotor and thus provides essentially rotating endplates to minimize the relative motion between the rotor/vane assembly and the endplates. This arrangement unnecessarily complicates the construction of the pump. FIGS. 27, 28(I), and 28(II) show the replacement of the annular roller bearings shown in FIGS. 1 to 3 by a solid annular ring that is equipped with various hydrodynamic grooves that purport to help the ring run on a lubricant film. However, the vane extensions rub against this rotating ring just as in the previously discussed arrangements. FIG. 29 shows another embodiment in which only a single outer annular ball bearing is utilized against which the vane extensions rub. In none of the disclosed embodiments, however, is any mechanism disclosed for greatly minimizing or totally eliminating this rubbing friction.
As will be seen in considerable detail hereinafter, the present invention not only eliminates the majority of the mechanical sliding friction endemic to previous techniques, but does so with fewer and simpler components than were required by the prior art. At the same time, my invention accomplishes the fundamentally important positive bi-axial radial vane motion control necessary for the practical operation of such machines. Finally, my invention accommodates the natural motion of the tips of circularly-tethered vanes by providing exceedingly close non-contact vane tip sealing as a result of properly shaping the mating or conjugate interior of the casing wall.
The embodiments shown and described herein are ideally suited for use as an automotive air conditioning compressor, although my invention may be used in many other applications and relationships. A major aspect of my present invention is comprised of embodiments of a freely rotating bearing (FRB), which center upon simple, anti-friction, easily-producible, economical, and positive retention or motion-positive means of ensuring the accurate transfer of radial movement from the circular radial vane guide to the vane. The cooperation of these means of precise anti-friction vane motion control with the proper internal casing profile, which I prefer to call a conjugate casing contour, results in maintaining an excellent sealing but non-contact and, thus, minimum friction relationship between the tips of the vanes and the internal conjugate stator contour. Such a condition yields a simple vane type fluid handling device of a high volumetric and energy efficiency.
An early actual prototype compressor of the type disclosed herein demonstrated in practice the efficacy of my invention. For example, while operating at only 1000 rpm with R-12 refrigerant at 40.degree. F. evaporation and 120.degree. F. condensation, the volumetric and adiabatic/isentropic efficiencies were measured respectively to be 81% and 88%. Conventional well-developed compressors operating under these demanding conditions produce volumetric and adiabatic/isentropic efficiencies on the order of only 60% and 70%, respectively.
One of the principal vane motion control embodiments involves the use of plain arc segment vane tethers that are pinned pivotally to the vanes and that ride directly upon freely-rotating retained rollers that roll inside the internal surface of the circular, non-rotating radial vane endplate guides. Another embodiment involves vane tether elements resembling roller skates, also pivotally-pinned to the vanes, that ride in non-rotating circular vane guides located in the endplates of the device.
The present invention fully eliminates all the components of rubbing between the pivoting vane gliders or tethers and the anti-friction rolling elements rolling within this annular channel race. This has been achieved by allowing full freedom of constrained rolling motion of the rollers contained within the outer channel race and the outer surface of the vane glider. However, for start-and-stop conditions, a simple retainer ring placed in the inside diameter region of the rollers is included to ensure that the rollers never fall out or get caught between respective vane gliders.
According to yet another embodiment of my invention, I provide another way of minimizing the friction between the vane gliders and the end-plate annuli. This embodiment is non-mechanical and depends upon the use of special low-friction interface materials or surface coatings. The use of such an annular liner is quite practical at least for moderate speeds. It is of commercial interest (even though its frictional component will necessarily be higher than pure rolling), especially because it is lower in cost compared to the true rolling element embodiment.
The present invention is distinguishable from the engine disclosed in the above-discussed U.S. Pat. No. 2,469,510 by virtue of the absence in my invention of the rigid rotating roller races of the conventional roller bearings that interface with the vane extensions. The elimination of these ground and hardened races by using a pivoting tether that is pinned to the vanes in lieu of the rigid vane extensions substantially reduces sliding friction and hence wear.
The bearing utilized in the present invention replaces the standard integral bearings comprising rigid inner and outer races separated by roller or ball rolling elements. Specifically, the single rigid inner race of the conventional type of roller bearing has been eliminated and replaced by separate individual sections in the form of the vane gliders that are supported directly by the anti-friction rolling elements with no intervening rigid race. Not only does this permit direct transmission of rolling motion, but the individual vane gliders independently accommodate the cyclic variation of circumferential velocity of the vanes. This constantly-changing circumferential velocity occurs, of course, due to the eccentricity of the rotor with respect to the stator and, again, is required in order to change the volume of the compressing gas pockets. The physically independent or separate nature of the rollers eliminate the sliding friction component inherent in prior art devices which relied upon a single rigid inner race.
A compressor built in accordance with the principles of my invention directly transfers the motion of varying speed vanes to essentially independent rolling elements. There is no intervening interface component such as the rotating ball bearing race as is shown in U.S. Pat. No. 2,469,570 that, due to its rigidity as a continuous ring, necessitates physical rubbing between the vane motion control extensions and the rotating rings. While those of a conventional "unit cage" made up of conventional rollers retained by conventional roller cages can be used, a slight component of rubbing will occur between the rollers and the cage. Even this component of rubbing can be eliminated, however, by providing a retainer cage for the rollers with sufficient space between rollers to allow the variable velocity rollers to freely rotate between the race annuli and the surfaces of the vane guide.
Similarly, none of the embodiments shown in U.S. Pat. No. 4,958,995 are true anti-friction devices in the sense of the present invention, notwithstanding that they seek to achieve anti-friction vane control. The prior art antifriction devices, unlike the present invention, produce large components of rubbing and sliding friction between the so-called anti-friction means and the vane extensions regardless of configuration because the vane extensions engage a rigidly rotating annular subcomponent. Moreover, the precise operating or conjugate shape which I have found to be required of the inner stator profile (or the vane tip profile) for non-contact vane machines has not been previously taught or achieved. My invention thus provides the very small physical clearances between the dynamic interfaces so as, on one hand, to minimize leakage and, on the other hand, to avoid physical contact which produces rubbing friction and wear.
I have also found it possible to achieve stator and vane tip circularization, i.e. a true circular interior profile and a true circular arc-shaped vane tip to achieve exact conjugate non-contact sealing geometry. The stator/vane tip circularization has been achieved by placing the center of the vane tip radius at the exact coincidence with the glider axle center and making the tip radius equal to the difference in the radii of the stator and the radius of the path of the glider axle centerline. The result is a truly circular sealing noncontact interior stator contour. Of course, this radius must be reduced by the amount of tip clearance desired, e.g. 0.001" to 0.002". Alternatively, this clearance can be achieved by increasing the radius of the circular stator by the same amount.
As will become more apparent as the detailed description proceeds, the anti-friction vane motion control embodiments revealed here are combinable to yield yet additional embodiments which can be used quite effectively, depending upon the purpose to be served.
It is therefore primary object of my invention to provide a vane type fluid displacement machine that accomplishes non-contact vane tip sealing in a particularly simple and energy efficient manner and which is relatively easy to manufacture and to maintain in service.
It is another important object of my invention to provide a non-contact rotary vane machine that is extremely reliable, and which can operate with a wide variety of refrigerants, including those not harmful to the Earth's stratospheric ozone layers.
It is still yet another object of my invention to provide a non-contact vane type compressor whose vane tips are positioned by the utilization of circular radial vane guides, eliminating the use of the costly non-circular vane guides extensively utilized by the prior art.